1. Field of the Invention
This invention relates generally to rotor devices and, more particularly to screw rotors.
2. Description of Related Art
Screw rotors are generally known to be used in compressors, expanders, and pumps. For each of these applications, a pair of screw rotors have helical threads and grooves that intermesh with each other in a housing. For an expander, a pressurized gaseous working fluid enters the rotors, expands into the volume as work is taken out from at least one of the rotors, and is discharged at a lower pressure. For a compressor, work is put into at least one of the rotors to compress the gaseous working fluid. Similarly, for a pump, work is put into at least one of the rotors to pump the liquid. The working fluid, either gas or liquid, enters through an inlet in the housing, is positively displaced within the housing as the rotors counter-rotate, and exits through an outlet in the housing.
The rotor profiles define sealing surfaces between the rotors themselves between the rotors and the housing, thereby sealing a volume for the working fluid in the housing. The profiles are traditionally designed to reduce leakage between the sealing surfaces, and special attention is given to the interface between the rotors where the threads and grooves of one rotor respectively intermesh with the grooves and threads of the other rotor. The meshing interface between rotors must be designed such that the threads do not lock-up in the grooves, and this has typically resulted in profile designs similar to gears, having radially widening grooves and tightly spaced involute threads around the circumference of the rotors. However, an involute for a gear tooth is primarily designed for strength and to prevent lock-up as teeth mesh with each other and are not necessarily optimum for the circumferential sealing of rotors within a housing.
The performance characteristics of screw rotors depend on several factors, including thermodynamic efficiencies, volumetric efficiencies, and mechanical efficiencies. Adiabatic efficiency is one type of parameter to evaluate the thermodynamic efficiency of a screw rotor system. Adiabatic efficiency is the ratio of the adiabatic horsepower required to compress a given amount of gas to the actual horsepower expended in the compressor cylinder. Volumetric efficiency is the ratio of the actual volume of working fluid flowing through the screw rotor, such as in one complete revolution, to the geometric volume of the screw rotor measured, which is also measured for one complete revolution. Mechanical efficiencies can include the efficiencies of any gear train that may be used to keep the rotors in proper phase with each other, bearings, and seals.
Although adiabatic efficiency and volumetric efficiency are different performance parameters, a number of screw rotor features can affect both of these efficiencies. For example, tightening tolerances between the rotors and the housing can improve both the volumetric efficiency and the adiabatic efficiency of a given rotor design. However, if tolerances are too tight for a given design, the volumetric efficiency may be improved while the adiabatic efficiency drops. Such performance characteristic could be caused by thermal expansion of the rotors, machining tolerances, and even the material properties of the rotors, which can result in intermittent contact between the rotors and the sides of the housing or between the rotors themselves.
Generally, one of the best ways to improve thermodynamic efficiencies is by keeping tight tolerances and minimizing leak pathways between the rotors and the housing and between the rotors themselves. However, in prior art screw rotors, leak pathways are inherent in the actual design of the rotors, i.e., the leaks can be reduced but not eliminated. Such inherent leaks would occur even when the tolerances are perfected, i.e., zero thermal expansion, perfect machining tolerances, and a perfectly smooth finished material. These leak pathways result in losses that adversely affect both the thermodynamic efficiency and the volumetric efficiency of screw rotors.
Accordingly, leak pathways are some of the most important losses to consider for the performance of screw rotors when the screw rotors are being designed because these losses negatively affect both thermodynamic efficiency and volumetric efficiency. Even with this knowledge that leak pathways should be minimized, the design methodology used for screw rotors produces these pathways as an inherent aspect of traditional screw rotor profiles. In fact, it is a common belief by the designers, manufacturers and users of screw rotors that it is impossible to eliminate some of the leaks in a screw rotor system. For example, according to Mattai Compressors, Inc., at its web site www.matteicomp.com/About/ScrewCompressors/, this belief is concisely stated even as this application is being filed in March 2004: “The technical problem is typical of the geometry of screw compressors. All screw manufacturers have tried to reduce the effect of the ‘blow hole’ by analyzing and adapting new rotor profiles to create smaller openings at the critical point, but its complete elimination is impossible.” Accordingly, to minimize the leak pathways, it is common knowledge that the rotors should seal perfectly along the contact line, but a number of prior art references also teach that the contact line should be as short as possible, i.e., should not extend to cusps on opposite sides of the housing. Several embodiments of short contact lines are set forth in the applicant's patent application Ser. No. 10/283,421 (Pub. No. 2003/0077198) and U.S. application Ser. No. 10/283,422. However, there remains a need for better methodologies for designing screw rotor profiles that account for machining constraints, thermal expansion and material tolerances, as well as mechanical efficiencies, and that also eliminate any inherent leak pathway from the design process, even though it is presently considered impossible. One example of a machining constraint set forth in the prior art is the need for blunt edges because of the concern that sharp edges have a tendency to break, e.g., U.S. Pat. No. 2,486,770.
Once the leak pathway problem is eliminated from the design methodology, i.e., screw rotor profiles that do inherently produce a leak pathway, the designer can balance all of the rotors' performance characteristics. For example, a rotor design without any inherent leak pathway may be slightly changed to include a small gap or leak pathway to permit another aspect to improve the rotors' overall performance at a given design point, i.e., tighter tolerances at steady state operation with thermal expansion. In comparison, when the leak pathway remains an inherent feature of the rotor profiles, the designer must first minimize the leak pathway using more complex designs that are harder and costlier to manufacture and then changes to the design are limited by the complexity of the design, machining and other manufacturing capabilities and thermal expansion requirements. Therefore, a new design methodology that produces screw rotor profile shapes without any leak pathways is needed. Additionally, it would also be advantageous if sharp-edged shapes that eliminate leak pathways and do not have a tendency to break could be designed and manufactured.
Leak pathways are generally caused by internal leakage between the rotors and the housing and between the rotors themselves and result in volumetric losses and thermodynamic losses due to recirculation of the working fluid within the rotors. For example, working fluid that is pressurized and leaks into a lower pressure region of the rotors is caused to expand to the lower pressure state with a higher temperature due to entropy and then must recirculate through the rotors before being expelled. Therefore, the overall temperature of entire rotor system, including the rotors and the working fluid, is increased due to the gain in entropy. Internal leakage is detected specifically at the following points:                (1) gaps between the inlet port and/or outlet port in the housing and the rotors, resulting in less than complete capture or ejection of the working fluid through the rotors;        (2) gaps between the outer periphery of each rotor and the inner surface of the housing, through which the working fluid leaks around the top land of a thread or the ridge of a groove to an adjacent working volume, respectively;        (3) gaps between the front and back of the intermeshing male rotor thread and female rotor groove, through which the working fluid leaks from the pressurized side to the suction side; and        (4) a gap formed on the front side of the rotor in the transition region, where the male rotor threads intermesh with the female rotor grooves proximate to the cusp of the cylindrical bores and which generally forms a tetrahedron (or a triangular shape in two-dimensions) that is defined by the shape of the gap between the intermeshing thread and groove and the cusp, and another gap similarly formed on the back side of the rotor, through which the working fluid leaks from one V-shaped working volume to an adjacent V-shaped working volume, i.e. commonly referred to as a blow hole, and through which the working fluid leaks from a pressurized region to a less pressurized region or to a suction region.        
As discussed above, threads must provide seals between the rotors and the walls of the housing and between the rotors themselves, and in all designs before the present invention, there has been a transition from sealing around the circumference of the housing to sealing between the rotors. In this transition, a gap is formed between the meshing threads and the housing, causing leaks of the working fluid through the gap in the sealing surfaces and resulting in less efficiency in the rotor system. A number of arcuate profile designs improve the seal between rotors and may reduce the gap in this transition region but these profiles still retain the characteristic gear profile with tightly spaced teeth around the circumference, resulting in a number of gaps in the transition region that are respectively produced by each of the threads. Some pumps minimize the number of threads and grooves and may only have a single acme thread for each of the rotors, but these threads have a wide profile around the circumferences of the rotors and generally result in larger gaps in the transition region.
Until now, screw rotor expanders, compressors and pumps have had similar fundamental flaws. Generally, they allow for leak pathways between the working side, i.e., expansion, compression or pumping, to the side that should be sealed from the working side for proper operation of the rotors, i.e., non-working. These rotor designs are commonly referred to as Roots-type rotors and Lysholm-type rotors. Krigar-type rotors, which are described in German Patent Nos. DE 4121 and DE 7116 from more than a century ago, have fallen out of favor, and this may possibly be due to the rise of the Lysholm-type rotors in the 1930's and 1940's. In an article entitled “A New Rotary Compressor” and written by Lysholm in the 1940's, Lysholm puts down the Krigar design as being unable to obtain any compression between the lobes with a two-thread/two-groove design (2×2 configuration). While it is clear from the images of the Krigar design that there definitely were sealing issues, especially between the threads and the grooves, and Krigar appears to be more directed to radial flow, the Lysholm conclusion that the Krigar design could not perform any compression with only the 2×2 configuration is flawed. Regardless, the industry and teachings have generally followed Lysholm and Roots with very little interest given to Krigar, except as a historical reference.
Based primarily on the Lysholm concept, many screw rotor designs have attempted to seal the male rotor with the female rotor and the housing, but the prior art designs have either a leak pathway between the rotors themselves or a leak pathway between the rotors and the housing, i.e., which according to the prior art quoted above, the elimination of which is “impossible.” In the past, the design of screw rotors have been based on profile designs that do not necessarily follow a mathematical formula, i.e., empirical design methodology, while other designs are based on particular curves or a combination of piecewise curves, i.e., formula design methodology, such as lines, arcs, circles, squares, trapezoids, involutes, inverse-involutes, parabolas, hyperbolas, cycloids, trochoids, epicycloids, epitrochoids, hypocycloids, hypotrochoids, as well as other straight and arcuate lines, and still other designs combine formula and empirical design methodologies. However, regardless of the design methodology, empirical or formula or a combination thereof, prior designs and respective methods for creating rotor profiles either explicitly teach or implicitly suggest and disclose creating the profile for the thread and corresponding groove using the shortest seal path between the rotors, i.e. the sealing region does not extend from the front cusp all the way to the back cusp. Additionally, many of the prior art methods are based on and remain similar to traditional gear design methods.
Some earlier designs have come close to a complete seal or may even be able to effect a complete seal in one pitch, see in particular co-pending U.S. application Ser. No. 10/283,422. Even for these single-pitch sealing rotors, some of the seals may only be along sealing lines, rather than sealing areas. Additionally, since the rotor profiles are designed according to the traditional gear profile design methods, these rotors are usually limited in the types of arcuate lines that can be used to effect the seal. Without accounting for the third dimension, the arcuate lines have typically been limited to epitrochoids, epicycloids, hypocycloids and other types of spirals, such as an Archimedean spiral.
When the third dimension is accounted for in prior art design methodologies, it is typically limited to standard helix angle definitions that have been developed for ordinary screws, i.e., fastening screws. Such an approach fails to truly account for and does not take advantage of the third dimension. It is well known that for any screw rotor, the helix angle of the grooves and threads vary depending on their depth. In particular, the top land of the thread has a lesser helix angle than the root of the thread, and the trough of the groove has a greater helix angle than the ridge of the groove. Accordingly, merely using a single helix angle for a rotor, such as the top land, the root, or any other single angle, even with a correction factor, has not accounted for the variations in the helix angles of the thread and the groove. In this way, the known screw rotor geometries are created using planar design methodologies for the rotor profiles rather than using a volumetric design methodology.
The planar design methodologies fail to apply the function of the helix angle with respect to the radius, resulting in the profiles with leak pathways discussed above. In one aspect, the planar design methods are unnecessarily restrictive because they only take advantage of two-dimensional space to overcome the limitation that the threads must not lock-up in the grooves. In another aspect, the planar design methods are not restrictive enough because when the profiles are expanded into three-dimensional space, the profiles have three-dimensional leak pathways. The extra degree of freedom provided by the third-dimension allows for a volumetric design that prevents lock-up while permitting perfect sealing between the male rotor and female rotor and between the rotors and the housing, a perfect seal which is equivalent to the complete seal of pistons. More generally, similar fundamental flaws in the prior art designs and their respective methodologies can be traced back to their failure to accommodate for and use the additional degree of design freedom provided by the third dimension. It is the additional degree of design freedom of volumetric design methodologies that permits an unlimited number of profile designs which effect a complete seal without locking up the rotors and without the unnecessary restrictions of the planar design methodologies.
For many prior art rotors, the leak pathway can be found between the face of the thread and the housing. In particular, the thread and groove are designed with significant curvatures at their top land edges and ridges according to the standard manner of designing meshing gear teeth. Such rounded edges and ridges cannot possibly seal between the rotors and the housing when the thread and groove begin meshing with each other. As the thread and groove rotate away from their seals with the housing and into their meshing positions with each other, the rounded edges produce a gap between the housing and the groove and/or the thread before the groove and thread actually mesh and reform a sealing line. The gap between the housing to groove and thread seal can be an order of magnitude greater than the tolerances for the seals between the between the rotors and the housing and the rotors themselves. In some designs, the gap can be even larger, such as in screw rotors that have a different number of threads and grooves, i.e. not the same number of threads as grooves, and the loss in pressure to the low pressure side causes the thermodynamic efficiency to drop. Therefore, the rotors must work harder to pump the same volume of air as compared with rotors according to the present invention which can maintain the same order of magnitude in the seal tolerances when each thread and respective groove begin meshing with each other as compared to the seal between the rotors and the housing and the rotors when in their fully intermeshed positions.
Additionally, by failing to take advantage of the third dimension in the design of the thread and groove, the prior art design methods have failed to optimize the basic screw rotor design or improve the screw rotor efficiencies to their full potential. As discussed above, the prior art design methodologies generally use planar coordinates to define the thread and groove profiles, and the third dimension is merely considered for the helix angle of the profiles. In an attempt to compensate for this unwitting failure to take advantage of the third dimension, the prior art designs have increasingly become more complex over the years without offering much improvement in the thermodynamic efficiency of the rotor system. As evidence of the failure to appreciate volumetric design methodologies as an alternative to traditional gear design methods combined with traditional fastener screw methods, these planar design methodologies increasingly led to these more complex screw rotor designs as machining and other manufacturing methods improved over the years and permitted the increasing complexity. Additionally, these increasingly complex screw rotor profile designs, which need such improved manufacturing methods, support the conclusion that the failure to take advantage of the third dimension has been an unwitting failure because volumetric design methodologies actually permit much more simplified designs which can be less complex to manufacture than profiles created using the planar design methodologies.